Auto rack rail road cars are used to transport automobiles. Typically, auto-rack rail road cars are loaded in the “circus loading” manner, by driving vehicles into the cars from one end, and securing them in place with chocks, chains or straps. When the trip is completed, the chocks are removed, and the cars are driven out. The development of autorack rail road cars can be traced back 80 or 90 years, when mass production led to a need to transport large numbers of automobiles from the factory to market.
Automobiles are a high value, relatively low density, relatively fragile type of lading. Damage to lading due to dynamic loading in the railcar may tend to arise principally in two ways. First, there are longitudinal input loads transmitted through the draft gear due to train line action or shunting. Second, there are vertical, rocking and transverse dynamic responses of the rail road car to track perturbations as transmitted through the rail car suspension. It would be desirable to improve ride quality to lessen the chance of damage occurring.
In the context of longitudinal train line action, damage most often occurs from two sources (a) slack run-in and run out; (b) humping or flat switching. Rail road car draft gear have been designed against slack run-out and slack run-in during train operation, and also against the impact as cars are coupled together. Historically, common types of draft gear, such as that complying with, for example, AAR specification M-901-G, have been rated to withstand an impact at 5 m.p.h. (8 km/h) at a coupler force of 500,000 Lbs. (roughly 2.2×106 N). Typically, these draft gear have a travel of 2¾ to 3 1/4  inches in buff before reaching the 500,000 Lbs. load, and before “going solid”. The term “going solid” refers to the point at which the draft gear exhibits a steep increase in resistance to further displacement. If the impact is large enough to make the draft gear “go solid” then the force transmitted, and the corresponding acceleration imposed on the lading, increases sharply. While this may be acceptable for ores, coal or grain, it is undesirably severe for more sensitive lading, such as automobiles or auto parts, rolls of paper, fresh fruit and vegetables and other high value consumer goods such as household appliances or electronic equipment. Consequently, from the relatively early days of the automobile industry there has been a history of development of longer travel draft gear to provide lading protection for relatively high value, low density lading, in particular automobiles and auto parts, but also farm machinery, or tractors, or highway trailers.
The subject of slack action is discussed at length in my co-pending U.S. patent application Ser. No. 09/920,437 filed Aug. 1, 2001 and incorporated herein by reference.
Since automobiles tend to be a relatively low density form of lading as compared to grain, ores, or coal, the volumetric capacity of the cars tends to be filled up before the weight of the reaches the maximum allowable weight for the trucks. This has led to efforts to increase the volumetric capacity of the cars. Over time, particularly in the period of 1945–1970, autotrack cars grew longer and taller. At present, an autorack car may be up to about 90 feet long and 20 ft-2 inches tall. Autorack cars may typically have a tall, somewhat barn-like housing. The housing has end doors that are intended to keep out thieves and vandals.
The desire to increase the internal volume of the autorack car, and the relatively light weight of the lading, led to the development of a special 70 Ton rail road car truck for use with autorack cars. A 70 Ton “special” truck is shown in the 1997 Car and Locomotive Cyclopedia (Simmons-Boardman, Omaha, 1997) at page 726. The illustration indicates that the total loading of the spring groups at solid is indicated as 70,166 Lbs. per spring group, giving a total of 140,334 Lbs. per truck and 280,668 Lbs. per single unit autorack car. The spring rate is indicated as 18,447 Lbs./in., per spring group or 36,894 Lbs./in for the truck overall (there being one spring group per side frame, and two spring groups per truck). The truck shown in the 1997 Cyclopedia is a swing motion truck manufactured by National Castings Inc. In contrast to a regular 70 Ton truck that has, typically, 33 inch diameter wheels, the 70 Ton special autorack truck has wheels that have a diameter of only 28 inches. This tends to allow for lower main deck wheel trackways, and hence greater inside clearance height. In part, the use of such a truck in an autorack car may reflect the low density of the lading. That is, a regular 70 Ton truck is designed to carry a gross weight on rail of 110,000 Lbs, for a total full car weight of 220,000 Lbs. If the dead sprung weight of a conventional single unit autorack car is 75–85,000 Lbs., and the unsprung weight is about 15,000 Lbs, that would leave about 120,000 Lbs., for lading. Assuming that a typical passenger sedan weighs about 2500 Lbs., that would allow for about 48 automobiles before the gross weight on rail would be exceeded. Even for larger, heavier vehicles, weighing perhaps as much as 5000 Lbs., this would still give some 24 light trucks, vans, or “sport utility vehicles”. But the volumetric capacity of a single unit autorack rail road car may be about 12–15 family sedans and perhaps fewer light trucks, vans, or SUV's. Thus the autorack rail road car truck loading may often tend to be significantly less than 110,000 lbs.
In contrast to the philosophy underlying the design of the 70 Ton special 28 inch truck, the present inventor believes that it is advantageous to use a truck having wheels larger than 33 inches in diameter for auto rack rail road cars. Wheel life and maintenance are dependent on wheel loading, and, for the same loading history, inversely dependent on wheel diameter. A larger wheel may tend to have lower operating stresses for the same lading; may tend to have a greater wear allowance for braking; may tend to undergo fewer rotations than a wheel of smaller diameter for the same distance travelled, and therefore may tend to accumulate fewer cycles in terms of fatigue life; and may tend not to get as hot during braking. All of these factors may tend to increase wheel life and reduce maintenance. Further, a larger wheel diameter may be used in conjunction with the use of longer springs. The use of longer springs may permit the employment of springs having a softer spring rate, giving a gentler ride. In terms of fatigue life and wear, this in turn may tend to give a load history with reduced peak loads, and lower frequency of those peak loads. Attainment of any one of these advantages would be desirable.
In terms of dynamic response through the trucks, there are a number of loading conditions to consider. First, there is a direct vertical response in the “vertical bounce” condition. This may typically arise when there is a track perturbation in both rails at the same point, such as at a level crossing or at a bridge or tunnel entrance where there may be a relatively sharp discontinuity in track stiffness. A second “rocking” loading condition occurs when there are alternating track perturbations, typically such as used formerly to occur with staggered spacing of 39 ft rails. This phenomenon is less frequent given the widespread use of continuously welded rails, and the generally lower speeds, and hence lower dynamic forces, used for the remaining non-welded track. A third loading condition arises from elevational changes between the tracks, such as when entering curves in which case a truck may have a tendency to warp. A fourth loading condition arises from truck “hunting”, typically at higher speeds, where the truck oscillates transversely between the rails. During hunting, the trucks tend most often to deform in a parallelogram manner. Fifth, lateral perturbations in the rails sometimes arise where the rails widen or narrow slightly, or one rail is more worn than another, and so on.
There are both geometric and historic factors to consider related to these loading conditions and the dynamic response of the truck. One historic factor is the near universal usage of the three-piece style of freight car truck in North America. While other types of truck are known, the three piece truck is overwhelmingly dominant in freight service in North America. The three piece truck relies on a primary suspension in the form of a set of springs trapped in a “basket” between the truck bolster and the side frames. Rather than requiring independent suspension of each wheel, for wheel load equalisation a three piece truck uses one set of springs, and the side frames pivot about the truck bolster ends in a manner like a walking beam. It is a remarkably simple and durable layout. However, the dynamic performance of the truck flows from that layout. The 1980 Car & Locomotive Cyclopedia, states at page 669 that the three piece truck offers “interchangeability, structural reliability and low first cost but does so at the price of mediocre ride quality and high cost in terms of car and track maintenance”. It would be desirable to retain many or all of these advantages while providing improved ride quality.
In terms of rail road car truck suspension loading regimes, the first consideration is the natural frequency of the vertical bounce response. The static deflection from light car (empty) to maximum laded gross weight (full) of a rail car at the coupler tends to be typically about 2 inches. In addition, rail road car suspensions have a dynamic range in operation, including a reserve travel allowance.
In typical historical use, springs were chosen to suit the deflection under load of a full coal car, or a full grain car, or fully loaded general purpose flat car. In each case, the design lading tended to be very heavy relative to the rail car weight. For example, the live load for a 286,000 lbs. car may be of the order of five times the weight of the dead sprung load (i.e., the weight of the car, including truck bolsters but less side frames, axles and wheels). Further, in these instances, the lading may not be particularly sensitive to abusive handling. That is, neither coal nor grain tends to be badly damaged by poor ride quality. As a result, these cars tend to have very stiff suspensions, with a dominant natural frequency in vertical bounce mode of about 2 Hz. when loaded, and about 4 to 6 Hz. when empty. Historically, much effort has been devoted to making freight cars light for at least two reasons. First, the weight to be back hauled empty is kept low, reducing the fuel cost of the backhaul. Second, as the ratio of lading to car weight increases, a higher proportion of hauling effort goes into hauling lading, rather than hauling the railcar.
By contrast, an autorack car, or other type of car for carrying relatively high value, low density lading such as auto parts, electronic consumer goods, or white goods more generally, has the opposite loading profile. A two unit articulated autorack car may have a light car (i.e., empty) weight of 165,000 lbs., and a lading weight when fully loaded of only 35–40,000 lbs., per car body unit. That is, not only may the weight of the lading be less than the sprung weight of the rail road car unit, it may be less than 40% of the car weight. The lading typically has a high, or very high, ratio of value to weight. Unlike coal or grain, automobiles are relatively fragile, and hence more sensitive to a gentle (or a not so gentle) ride. As a relatively fragile, high value, high revenue form of lading, it may be desirable to obtain superior ride quality to that suitable for coal or grain.
Historically, auto rack cars were made by building a rack structure on top of a general purpose flat car. As such, the resultant car was sprung for the flat car design loads. When loaded with automobiles, this might yield a vertical bounce natural frequency in the range of 3 Hz. It would be preferable for the railcar vertical bounce natural frequency to be on the order of 1.4 Hz or less when loaded. Since this natural frequency varies as the square root of the quotient obtained by dividing the spring rate of the suspension by the overall sprung mass, it is desirable to reduce the spring constant, to increase the mass, or both.
One way to improve ride quality is to increase the dead sprung weight of the rail road car body. Deliberately increasing the mass of a freight car is counter intuitive, since many years of effort has gone into reducing the weight of rail cars relative to the weight of the lading for the reasons noted above. One manufacturer, for example, advertises a light weight aluminium auto-rack car. However, given the high value and low density of the lading, adding weight may be reasonable to obtain a desired level of ride quality. Further, auto rack rail cars tend to be tall, long, and thin, with the upper deck loads carried at a relatively high location as measured from top of rail. A significant addition of weight at a low height relative to top of rail may also be beneficial in reducing the height of the center of gravity of the loaded car.
Another way to improve ride quality is to decrease the spring rate. Decreasing the spring rate involves further considerations. Historically the deck height of a flat car tended to be very closely related to the height of the upper flange of the center sill. This height was itself established by the height of the cap of the draft pocket. The size of the draft pocket was standardised on the basis of the coupler chosen, and the allowable heights for the coupler knuckle. The deck height usually worked out to about 41 inches above top of rail. For some time auto rack cars were designed to a 19 ft height limit. To maximise the internal loading space, it has been considered desirable to lower the main deck as far as possible, particularly in tri-level cars. Since the lading is relatively light, the rail car trucks have tended to be light as well, such as 70 Ton trucks, as opposed to 100, 110 or 125 Ton trucks for coal, ore, or grain cars at 263,000, 286,000 or 315,000 lbs. gross weight on rail. Since the American Association of Railroads (AAR) specifies a minimum clearance of 5″ above the wheels, the combination of low deck height, deck clearance, and minimum wheel height set an effective upper limit on the spring travel, and reserve spring travel range available. If softer springs are used, the remaining room for spring travel below the decks may well not be sufficient to provide the desired reserve height. In consequence, the present inventor proposes, contrary to lowering the main deck, that the main deck be higher than 42 inches to allow for more spring travel.
As noted above, many previous auto rack cars have been built to a 19 ft height. Another major trend in recent years has been the advent of “double stack” intermodal container cars capable of carrying two shipping containers stacked one above the other in a well or to other freight cars falling within the 20 ft 2 in. height limit of AAR plate H. Many main lines have track clearance profiles that can accommodate double stack cars. Consequently, it is now possible to use auto rack cars built to the higher profile of the double stack intermodal container cars.
While decreasing the primary vertical bounce natural frequency appears to be advantageous for auto rack rail road cars generally, including single car unit auto rack rail road cars, articulated auto rack cars may also benefit not only from adding ballast, but from adding ballast preferentially to the end units near the coupler end trucks. As explained more fully in the description below, the interior trucks of articulated cars tend to be more heavily burdened than the end trucks, primarily because the interior trucks share loads from two adjacent car units, while the coupler end trucks only carry loads from one end of one car unit. It would be advantageous to even out this loading so that the trucks have roughly similar vertical bounce frequencies.
Three piece trucks currently in use tend to use friction dampers, sometimes assisted by hydraulic dampers such as can be mounted, for example, in the spring set. Friction damping has most typically been provided by using spring loaded blocks, or snubbers, mounted with the spring set, with the friction surface bearing against a mating friction surface of the columns of the side frames, or, if the snubber is mounted to the side frame, then the friction surface is mounted on the face of the truck bolster. There are a number of ways to do this. In some instances, as shown at p. 847 of the 1961 Car Builders Cyclopedia lateral springs are housed in the end of the truck bolster, the lateral springs pushing horizontally outward on steel shoes that bear on the vertical faces of the side columns of the side frames. This provides roughly constant friction (subject to the wear of the friction faces), without regard to the degree of compression of the main springs of the suspension.
In another approach, as shown at p. 715 of the 1997 Car & Locomotive Cyclopedia, one of the forward springs in the main spring group, and one of the rearward springs in the main spring group bear upon the underside, or short side, of a wedge. One of the long sides, typically an hypotenuse of a wedge, engages a notch, or seat, formed near the outboard end of the truck bolster, and the third side has the friction face that abuts, and bears against, the friction face of the side column (either front or rear, as the case may be), of the side frame. The action of this pair of wedges then provides damping of the various truck motions. In this type of truck the friction force varies directly with the compression of the springs, and increases and decreases as the truck flexes. In the vertical bounce condition, both friction surfaces work in the same direction. In the warping direction (when one wheel rises or falls relative to the other wheel on the same side, thus causing the side frame to pivot about the truck bolster) the friction wedges work in opposite directions against the restoring force of the springs.
The “hunting” phenomenon has been noted above. Hunting generally occurs on tangent (i.e., straight) track as railcar speed increases. It is desirable for the hunting threshold to occur at a speed that is above the operating speed range of the rail car. During hunting the side frames tend to want to rotate about a vertical axis, to a non-perpendicular angular orientation relative to the truck bolster sometimes called “parallelogramming” or lozenging. This will tend to cause angular deflection of the spring group, and will tend to generate a squeezing force on opposite diagonal sides of the wedges, causing them to tend to bear against the side frame columns. This diagonal action will tend to generate a restoring moment working against the angular deflection. The moment arm of this restoring force is proportional to half the width of the wedge, since half of the friction plate lies to either side of the centreline of the side frame. This tends to be a relatively weak moment connection, and the wedge, even if wider than normal, tends to be positioned over a single spring in the spring group.
Typically, for a truck of fixed wheelbase length, there is a trade-off between wheel load equalisation and resistance to hunting. Where a car is used for carrying high density commodities at low speeds, there may tend to be a higher emphasis on maintaining wheel load equalisation. Where a car is light, and operates at high speed there will be a greater emphasis on avoiding hunting. In general, the parallelogram deformation of the truck in hunting may be deterred by making the truck laterally more stiff. One approach to discouraging hunting is to use a transom, typically in the form of a channel running from between the side frames below the spring baskets. Another approach is to use a frame brace.
One way to address the hunting issue is to employ a truck having a longer wheelbase, or one whose length is proportionately great relative to its width. For example, at present two axle truck wheelbases may range from about 5′-3″ to 6′-0″. However, the standard North American track gauge is 4′–8½″, giving a wheelbase to track width ratio possibly as small as 1.12. At 6′-0″ the ratio is roughly 1.27. It would be preferable to employ a wheelbase having a longer aspect ratio relative to the track gauge. As described herein, one aspect of the present invention employs a truck with a longer wheelbase, which may be about 80 to 86 inches, giving a ratio of 1.42 or 1.52. This increase in wheelbase length may tend also to be benign in terms of wheel loading equalisation.
In a typical spring seat and spring group arrangement, the side frame window may typically be of the order of 21 inches in height from the spring seat base to the underside of the overarching compression member, and the width of the side frame window between the wear plates on the side frame columns is typically about 18″, giving a side frame window that is taller than wide in the ratio of about 7:6. Similarly, the bottom spring seat has a base that is typically about 18 inches long to correspond to the width of the side frame window, and about 16 inches wide in the transverse direction, that is being longer than wide. It may be advantageous to make the side frame windows wider, and the spring seat correspondingly longer to accommodate larger diameter long travel springs with a softer spring rate or a larger number of softer coils of smaller diameter. At the same time, lengthening the wheel base of the truck may also be advantageous since it is thought that a longer wheelbase may ameliorate truck hunting performance, as noted above. Such a design change is counter-intuitive since it may generally be desired to keep truck size small, and widening the unsupported window span may not have been considered desirable heretofore.
Another way to raise the hunting threshold is to increase the parallelogram stiffness between the bolster and the side frames. It is possible, as described herein, to employ pairs of damper wedges, of comparable size to those previously used, the two wedges being placed side by side and each individually supported by a different spring, or being the outer two wedges in a three deep spring group, to give a larger moment arm to the restoring force and to the damping associated with that force.
One determinant of overall ride quality is the dynamic response to lateral perturbations. That is, when there is a lateral perturbation at track level, the rigid steel wheelsets of the truck may be pushed sideways relative to the car body. Lateral perturbations may arise for example from uneven track, or from passing over switches or from turnouts and other track geometry perturbations. When the train is moving at speed, the time duration of the input pulse due to the perturbation may be very short.
The suspension system of the truck reacts to the lateral perturbation. It is generally desirable for the force transmission to be relatively low. High force transmissibility, and corresponding high lateral acceleration, may tend not to be advantageous for the lading. This is particularly so if the lading includes relatively fragile goods, such as automobiles, electronic equipment, white goods, and other consumer products. In general, the lateral stiffness of the suspension reflects the combined displacement of (a) the sideframe between (i) the pedestal bearing adapter and (ii) the bottom spring seat (that is, the sideframes swing laterally as a pendulum with the pedestal bearing adapter being the top pivot point for the pendulum); and (b) the lateral deflection of the springs between (i) the lower spring seat in the sideframe and (ii) the upper spring mounting against the underside of the truck bolster, and (c) the moment and the associated angular displacement between the (i) spring seat in the sideframe and (ii) the upper spring mounting against the underside of the truck bolster.
In a conventional rail road car truck, the lateral stiffness of the spring groups is sometimes estimated as being approximately ½ of the vertical spring stiffness. Thus the choice of vertical spring stiffness may strongly affect the lateral stiffness of the suspension. The vertical stiffness of the spring groups may tend to yield a vertical deflection at the releasable coupler from the light car (i.e., empty) condition to the fully laden condition of about 2 inches. For a conventional grain or coal car subject to a 286,000 lbs., gross weight on rail limit, this may imply a dead sprung load of some 50,000 lbs., and a live sprung load of some 220,000 lbs., yielding a spring stiffness of 25–30,000 lbs./in., per spring group (there being, typically, two groups per truck, and two trucks per car). This may yield a lateral spring stiffness of 13–16,000 lbs./in per spring group. It should be noted that the numerical values given in this background discussion are approximations of ranges of values, and are provided for the purposes of general order-of-magnitude comparison, rather than as values of a specific truck.
The second component of stiffness relates to the lateral deflection of the sideframe itself. In a conventional truck, the weight of the sprung load can be idealized as a point load applied at the center of the bottom spring seat. That load is carried by the sideframe to the pedestal seat mounted on the bearing adapter. The vertical height difference between these two points may be in the range of perhaps 12 to 18 inches, depending on wheel size and sideframe geometry. For the general purposes of this description, for a truck having 36 inch wheels, 15 inches (±) might be taken as a roughly representative height.
The pedestal seat may typically have a flat surface that bears on an upwardly crowned surface of the bearing adapter. The crown may typically have a radius of curvature of about 60 inches, with the center of curvature lying below the surface (i.e., the surface is concave downward).
When a lateral shear force is imposed on the springs, there is a reaction force in the bottom spring seat that will tend to deflect the sideframe, somewhat like a pendulum. When the sideframe takes on an angular deflection in one direction, the line of contact of the flat surface of the pedestal seat with the crowned surface of the bearing adapter will tend to move along the arc of the crown in the opposite direction. That is, if the bottom spring seat moves outboard, the line of contact will tend to move inboard. This motion is resisted by a moment couple due to the sprung weight of the car on the bottom spring seat, acting on a moment arm between (a) the line of action of gravity at the spring seat and (b) the line of contact of the crown of the bearing adapter. For a 286,000 lbs. car the apparent stiffness of the sideframe may be of the order of 18,000–25,000 lbs./in, measured at the bottom spring seat. That is, the lateral stiffness of the sideframe (i.e., the pendulum action by itself) can be greater than the (already relatively high) lateral stiffness of the spring group in shear, and this apparent stiffness is proportional to the total sprung weight of the rail car (including lading). When taken as being analogous to two springs in series, the overall equivalent lateral spring stiffness may be of the order of 8,000 lbs./in. to 10,000, per sideframe. A car designed for lesser weights may have softer apparent stiffness. This level of stiffness may not always yield as smooth a ride as may be desired.
There is another component of spring stiffness due to the unequal compression of the inside and outside portions of the spring group as the bottom spring seat rotates relative to the upper spring group mount under the bolster. This stiffness, which is additive to (that is, in parallel with) the stiffness of the sideframe, can be significant, and may be of the order of 3000–3500 lbs./in per spring group, depending on the stiffness of the springs and the layout of the group. Other second and third order effects are neglected for the purpose of this description. The total lateral stiffness for one sideframe, including the spring stiffness, the pendulum stiffness and the spring moment stiffness, for a S2HD 110 Ton truck may be about 9200 lbs/inch per side frame.
It has been observed that it may be preferable to have springs of a given vertical stiffness to give certain vertical ride characteristics, and a different characteristic for lateral perturbations. In particular, a softer lateral response may be desired at high speed (greater than about 50 m.p.h) and relatively low amplitude to address a truck hunting concern, while a different spring characteristic may be desirable to address a low speed (roughly 10–25 m.p.h) roll characteristic, particularly since the overall suspension system may have a roll mode resonance lying in the low speed regime.
An alternate type of three piece truck is the “swing motion” truck. One example of a swing motion truck is shown at page 716 in the 1980 Car and Locomotive Cyclopedia (1980, Simmons-Boardman, Omaha). This illustration, with captions removed, is the basis of FIGS. 1a, 1b and 1c, herein, labelled “Prior Art”. Since the truck has both lateral and longitudinal axes of symmetry, the artist has only shown half portions of the major components of the truck. The particular example illustrated is a swing motion truck produced by National Castings Inc., more commonly referred to as “NACO”. Another example of a NACO Swing Motion truck is shown at page 726 of the 1997 Car and Locomotive Cyclopedia (1997, Simmons-Boardroom, Omaha). An earlier swing motion three piece truck is shown and described in U.S. Pat. No. 3,670,660 of Weber et al., issued Jun. 20, 1972, the specification of which is incorporated herein by reference.
In a swing motion truck, the sideframe is mounted as a “swing hanger” and acts much like a pendulum. In contrast to the truck described above, the bearing adapter has an upwardly concave rocker bearing surface, having a radius of curvature of perhaps 10 inches and a center of curvature lying above the bearing adapter. A pedestal rocker seat nests in the upwardly concave surface, and has itself an upwardly concave surface that engages the rocker bearing surface. The pedestal rocker seat has a radius of curvature of perhaps 5 inches, again with the center of curvature lying upwardly of the rocker.
In this instance, the rocker seat is in dynamic rolling contact with the surface of the bearing adapter. The upper rocker assembly tends to act more like a hinge than the shallow crown of the bearing adapter described above. As such, the pendulum may tend to have a softer, perhaps much softer, response than the analogous conventional sideframe. Depending on the geometry of the rocker, this may yield a sideframe resistance to lateral deflection in the order of ¼ (or less) to about ½ of what might otherwise be typical. If combined in series with the spring group stiffness, it can be seen that the relative softness of the pendulum may tend to become the dominant factor. To some extent then, the lateral stiffness of the truck becomes less strongly dependent on the chosen vertical stiffness of the spring groups at least for small displacements. Furthermore, by providing a rocking lower spring seat, the swing motion truck may tend to reduce, or eliminate, the component of lateral stiffness that may tend to arise because of unequal compression of the inboard and outboard members of the spring groups when the sideframe has an angular displacement, thus further softening the lateral response.
In the truck of U.S. Pat. No. 3,670,660 the rocking of the lower spring seat is limited to a range of about 3 degrees to either side of center, and a transom extends between the sideframes, forming a rigid, unsprung, lateral connecting member between the rocker plates of the two sideframes. In this context, “unsprung” refers to the transom being mounted to a portion of the truck that is not resiliently isolated from the rails by the main spring groups.
When the three degree condition is reached, the rockers “lock-up” against the side frames, and the dominant lateral displacement characteristic is that of the main spring groups in shear, as illustrated and described by Weber. The lateral, unsprung, sideframe connecting member, namely the transom, has a stop that engages a downwardly extending abutment on the bolster to limit lateral travel of the bolster relative to the sideframes. This use of a lateral connecting member is shown and described in U.S. Pat. No. 3,461,814 of Weber, issued Mar. 7, 1967, also incorporated herein by reference. As noted in U.S. Pat. No. 3,670,660 the use of a spring plank had been known, and the use of an abutment at the level of the spring plank tended to permit the end of travel reaction to the truck bolster to be transmitted from the sideframes at a relatively low height, yielding a lower overturning moment on the wheels than if the end-of-travel force were transmitted through gibs on the truck bolster from the sideframe columns at a relatively greater height. The use of a spring plank in this way was considered advantageous.
In Canadian Patent 2,090,031, (issued Apr. 15, 1997 to Weber et al.,) noting the advent of lighter weight, low deck cars, Weber et al., replaced the transom with a lateral rod assembly to provide a rigid, unsprung connection member between the platforms of the rockers of the lower spring seats. As noted above, one type of car in which relative lightness and a low main deck has tended to be found is an Autorack car.
For the purposes of rapid estimation of truck lateral stiffness, the following formula can be used:ktruck=2×[(ksideframe)−1+(kspring shear)31 1]−1
where                ksideframe=[kpendulum+kspring moment]        kspring shear=The lateral spring constant for the spring group in shear.        kpendulum=The force required to deflect the pendulum per unit of deflection, as measured at the center of the bottom spring seat.        kspring moment=The force required to deflect the bottom spring seat per unit of sideways deflection against the twisting moment caused by the unequal compression of the inboard and outboard springs.        
For the range of motion that may typically be of interest, and for small angles of deflection, kpendulum can be taken as being approximately constant at, for example, the value obtained for deflection of one degree. This may tend to be a sufficiently accurate approximation for the purposes of general calculation.
In a pure pendulum, the lateral constant for small angles approximates k=W/L, where k is the lateral constant, W is the weight, and L is the pendulum length. Further, for the purpose of rapid comparison of the lateral swinging of the sideframes, an equivalent pendulum length for small angles of deflection can be defined as Leq=W/kpendulum. In this equation W represents the sprung weight borne by that sideframe, typically ¼ of the total sprung weight for a symmetrical single unit rail car. For a conventional truck Leq may be of the order of about 3 or 4 inches. For a swing motion truck, Leq may be of the order of about 10 to 15 inches.
It is also possible to define the pendulum lateral stiffness (for small angles) in terms of the length of the pendulum, the radius of curvature of the rocker, and the design weight carried by the pendulum according to the formula:kpendulum=(Flateral/δlateral)=(W/Lpendulum)[(Rcurvature/Lpendulum)+1]
where:
kpendulum=the lateral stiffness of the pendulum
Flateral=the force per unit of lateral deflection
δlateral=a unit of lateral deflection
W=the weight borne by the pendulum
Lpendulum=the length of the pendulum, being the vertical distance from the contact surface of the bearing adapter to the bottom spring seat
Rcurvature=the radius of curvature of the rocker surface
Following from this, if the pendulum stiffness is taken in series with the lateral spring stiffness, then the resultant overall lateral stiffness can be obtained. Using this number in the denominator, and the design weight in the numerator yields a length, effectively equivalent to a pendulum length if the entire lateral stiffness came from an equivalent pendulum according toLresultant=W/k·lateral total
For a conventional truck with a 60 inch radius of curvature rocker, and stiff suspension, this length, Lresultant may be of the order of 6–8 inches, or thereabout.
So that the present invention may better be understood by comparison, in the prior art illustration of FIGS. 1a, 1b, and 1c, a NACO swing motion truck is identified generally as A20. In as much as the truck is symmetrical about the truck center both from side-to-side and lengthwise, the artist has shown only half of the bolster, identified as A22, and half of one of the sideframes, identified as A24.
In the customary manner, sideframe A24 has defined in it a generally rectangular window A26 that admits one of the ends of the bolster A28. The top boundary of window A26 is defined by the sideframe arch, or compression member identified as top chord member A30, and the bottom of window A26 is defined by a tension member, identified as bottom chord A32. The fore and aft vertical sides of window A26 are defined by sideframe columns A34.
At the swept up ends of sideframe A24 there are sideframe pedestal fittings A38 which each accommodate an upper rocker identified as a pedestal rocker seat A40, that engages the upper surface of a bearing adapter A42. Bearing adapter A42 itself engages a bearing mounted on one of the axles of the truck adjacent one of the wheels. A rocker seat A40 is located in each of the fore and aft pedestals, the rocker seats being longitudinally aligned such that the sideframe can swing transversely relative to the rolling direction of the truck A20 generally in what is referred to as a “swing hanger” arrangement.
The bottom chord of the sideframe includes pockets A44 in which a pair of fore and aft lower rocker bearing seats A46 are mounted. The lower rocker seat A48 has a pair of rounded, tapered ends or trunnions A50 that sit in the lower rocker bearings A48, and a medial platform A52. An array of four corner bosses A54 extend upwardly from platform A52.
An unsprung, lateral, rigid connecting member in the nature of a spring plank, or transom A60 extends cross-wise between the sideframes in a spaced apart, underslung, relationship below truck bolster A22. Transom A60 has an end portion that has an array of four apertures A62 that pick up on bosses A54. A grouping, or set of springs A64 seats on the end of the transom, the corner springs of the set locating above bosses A54.
The spring group, or set A64, is captured between the distal end of bolster A22 and the end portion of transom A60. Spring set A64 is placed under compression by the weight of the rail car body and lading that bears upon bolster A22 from above. In consequence of this loading, the end portion of transom A60, and hence the spring set, are carried by platform A54. The reaction force in the springs has a load path that is carried through the bottom rocker A70 (made up of trunnions A50 and lower rocker bearings A48) and into the sideframe A22 more generally.
Friction damping is provided by damping wedges A72 that seat in mating bolster pockets A74. Bolster pockets A74 have inclined damper seats A76. The vertical sliding faces of the friction damper wedges then ride up an down on friction wear plates A80 mounted to the inwardly facing surfaces of the sideframe columns.
The “swing motion” truck gets its name from the swinging motion of the sideframe on the upper rockers when a lateral track perturbation is imposed on the wheels. The reaction of the sideframes is to swing, rather like pendula, on the upper rockers. When this occurs, the transom and the truck bolster tend to shift sideways, with the bottom spring seat platform rotating on the lower rocker.
The upper rockers are inserts, typically of a hardened material, whose rocking, or engaging, surface A80 has a radius of curvature of about 5 inches, with the center of curvature when assembled) lying above the upper rockers (i.e., the surface is upwardly concave).
As noted above, one of the features of a swing motion truck is that while it may be quite stiff vertically, and while it may be resistant to parallelogram deformation because of the unsprung lateral connection member, it may at the same time tend to be laterally relatively soft.
The use of multiple variable friction force dampers in which the wedges are mounted over members of the spring group, is shown in U.S. Pat. No. 3,714,905 of Barber, issued Feb. 6, 1973. The damper arrangement shown by Barber is not apparently presently available in the market, and does not seem ever to have been made available commercially.
Notably, the damper wedges shown in Barber appear to have relatively sharply angled wedges, with an included angle between the friction face (i.e., the face bearing against the side frame column) and the sliding face (i.e., the angled face seated in the damper pocket formed in the bolster, typically the hypotenuse) of roughly 35 degrees. The angle of the third, or opposite, horizontal side face, namely the face that seats on top of the vertically oriented spring, is the complementary angle, in this example, being about 55 degrees. It should be noted that as the angle of the wedge becomes more acute, (i.e., decreasing from about 35 degrees) the wedge may have an undesirable tendency to jam in the pocket, rather than slide.
Barber, above, shows a spring group of variously sized coils with four relatively small corner coils loading the four relatively sharp angled dampers. From the relative sizes of the springs illustrated, it appears that Barber was contemplating a spring group of relatively traditional capacity—a load of about 80,000 lbs., at a “solid” condition of 3 1/16 inches of travel, for example, and an overall spring rate for the group of about 25,000 lbs/inch, to give 2 inches of overall rail car static deflection for about 200,000 lbs live load.
Apparently keeping roughly the same relative amount of damping overall as for a single damper, Barber appears to employ individual B331 coils (k=538 lb/in, (±)) under each friction damper, rather than a B432 coil (k=1030 lb/in, (±)) as might typically have been used under a single damper for a spring group of the same capacity. As such, it appears that Barber contemplated that springs accounting for somewhat less than 15% of the overall spring group stiffness would underlie the dampers.
These spring stiffness might typically be suitable for a rail road car carrying iron ore, grain or coal, where the lading is not overly fragile, and the design ratio of live load to dead sprung load is typically greater than 3:1. It might not be advantageous for a rail road car for transporting automobiles, auto parts, consumer electronics or other white goods of relatively low density and high value where the design ratio of live load to dead sprung load may be well less than 2:1, and quite possibly lying in the range of 0.4:1 to 1:1.
In the past, spring groups have been arranged such that the spring loading under the dampers has been proportionately small. That is, the dampers have typically been seated on side spring coils, as shown in the AAR standard spring groupings shown in the 1997 Car & Locomotive Cyclopedia at pages 743–746, in which the side spring coils, inner and outer as may be, are often B321, B331, B421, B422, B432, or B433 springs as compared to the main spring coils, such that the springs under the dampers have lower spring rates than the other coil combinations in the other positions in the spring group. As such, the dampers may be driven by less than 15% of the total spring stiffness of the group generally.
In U.S. Pat. No. 5,046,431 of Wagner, issued Sep. 10, 1991, the standard inboard-and-outboard gib arrangement on the truck bolster was replaced by a single central gib mounted on the side frame column for engaging the shoulders of a vertical channel defined in the end of the truck bolster. In doing this, the damper was split into inboard and outboard portions, and, further, the inboard and outboard portions, rather than lying in a common transverse vertical plane, were angled in an outwardly splayed orientation.
Wagner's gib and damper arrangement may not necessarily be desirable in obtaining a desired level of ride quality. In obtaining a soft ride it may be desirable that the truck be relatively soft not only in the vertical bounce direction, but also in the transverse direction, such that lateral track perturbations can be taken up in the suspension, rather than be transmitted to the car body, (and hence to the lading), as may tend undesirably to happen when the gibs bottom out (i.e., come into hard abutting contact with the side frame) at the limit of horizontal travel.
The present inventor has found it desirable that there be an allowance for lateral travel of the truck bolster relative to the wheels of the order of 1 to 1½ inches to either side of a neutral central position. Wagner does not appear to have been concerned with this issue. On the contrary, Wagner appears to show quite a tight gib clearance, with relatively little travel before solid contact. Furthermore, transverse displacement of the truck bolster relative to the side frame is typically resiliently resisted by the horizontal shear in the spring groups, and by the pendulum motion of the side frames rocking on the crowns of the bearing adapters, these two components being combined like springs in series. Wagner's canted dampers appear to make lateral translation of the bolster stiffer, rather than softer. This may not be advantageous for relatively fragile lading. In the view of the present inventor, while it is advantageous to increase resistance to the hunting phenomenon, it may not be advantageous to do so at the expense of increasing lateral stiffness.
In the damper groups themselves, it is thought that parallelogram deflection of the truck such that the truck bolster is not perpendicular to the side frame, as during hunting, may tend to cause the dampers to try to twist angularly in the damper seats. In that situation one corner of the damper may tend to be squeezed more tightly than the other. As a result, the tighter corner may try to retract relative to the less tight corner, causing the damper wedge to squirm and rotate somewhat in the pocket. This tendency to twist may also tend to reduce the squaring, or restoring force that tends to move the truck back into a condition in which the truck bolster is square relative to the side frames.
Consequently, it may be desirable to discourage this twisting motion by limiting the freedom to twist, as, for example, by introducing a groove or ridge, or keyway, or channel feature to govern the operation of the spring in the damper pocket. It may also be advantageous to use a split wedge to discourage twisting, such that one portion of the wedge can move relative to the other, thus finding a different position in a linear sense without necessarily forcing the other portion to twist. Further still, it may be advantageous to employ a means for encouraging a laterally inboard portion of the damper, or damper group, to be biased to its most laterally inboard position, and a laterally outboard portion of the damper, or the damper group, to be biased to its most laterally outboard position. In that way, the moment arm of the restoring force may tend to remain closer to its largest value. One way to do this, as described in the description of the invention, below, is to add a secondary angle to the wedge.
In the terminology herein, wedges have a primary angle ψ, namely the included angle between (a) the sloped damper pocket face mounted to the truck bolster, and (b) the side frame column face, as seen looking from the end of the bolster toward the truck center. This is the included angle described above. A secondary angle is defined in the plane of angle ψ, namely a plane perpendicular to the vertical longitudinal plane of the (undeflected) side frame, tilted from the vertical at the primary angle. That is, this plane is parallel to the (undeflected) long axis of the truck bolster, and taken as if sighting along the back side (hypotenuse) of the damper.
The secondary angle β is defined as the lateral rake angle seen when looking at the damper parallel to the plane of angle ψ. As the suspension works in response to track perturbations, the wedge forces acting on the secondary angle will tend to urge the damper either inboard or outboard according to the angle chosen. In as much as the tapered region of the wedge may be quite thin in terms of vertical through-thickness, it may be desirable to step the sliding face of the wedge (and the co-operating face of the bolster seat) into two or more portions. This may be particularly so if the angle of the wedge is large.
Split wedges and two part wedges having a chevron, or chevron like, profile when seen in the view of the secondary angle can be used. Historically, split wedges have been deployed as a pair over a single spring, the split tending to permit the wedges to seat better, and to remain better seated, under twisting condition than might otherwise be the case. The chevron profile of a solid wedge may tend to have the same intent of preventing rotation of the sliding face of the wedge relative to the bolster in the plane of the primary angle of the wedge. Split wedges and compound profile wedges can be employed in pairs as described herein.
In a further variation, where a single broad wedge is used, with a compound or other profile, it may be desirable to seat the wedge on two or more springs in an inboard-and-outboard orientation to create a restoring moment such as might not tend to be achieved by a single spring alone. That is, even if a single large wedge is used, the use of two, spaced apart springs may tend to generate a restoring moment if the wedge tries to twist, since the deflection of one spring may then be greater that the other.
When the dampers are placed in pairs, either immediately side-by-side or with spacing between the pairs, the restoring moment for squaring the truck will tend not only to be due to the increase in compression to one set of springs due to the extra tendency to squeeze the dampers downward in the pocket, but due to the difference in compression between the springs that react to the extra squeezing of one diagonal set of dampers and the springs that act against the opposite diagonal pair that will tend to be less tightly squeezed.